In a device of the type mentioned at the outset disclosed in DE 37 16 210 A1 the spring bars are arranged in planes perpendicular to the axis of rotation. In this case, the rotor can oscillate in a plane perpendicular to the spring bars and thus also effect pitch oscillations.
For balancing, articulated shafts are conventionally fixed horizontally at their ends to two rotatable spindles. Imbalance is measured at comparatively high rotational speeds, conventionally up to operating speed. The rotational speed measurements are predetermined and are adapted to the subsequent fitting conditions as the state of imbalance of articulated shafts can change depending on the rotational speed. This can be caused by play and friction in the joints and possibly on a sliding part and by curving of the articulated shaft housing.
In order to balance articulated shafts as close to operational conditions as possible, automatic articulated shaft balancing machines are required, in which the resonant masses and the rigidity of the mounting spindle support constitute conditions similar to operational conditions. A particularly simple embodiment of the resonant mass is deliberately avoided. In order to balance the articulated shaft as close to operational conditions as possible, the dynamic properties of the motor unit and differential gear, to which the articulated shaft is to be coupled, are reproduced in a configuration of the frame upper part, which configuration is of the aforementioned type and is substantially stringently conditional on isotropy and mass. The spring support of the oscillating frame upper part is thus also designed in such a way that, to measure the imbalance, the rotational speed lies above the resonant frequency of the bearing frame and the machine can therefore be operated in the “supercritical” region.
In a known articulated shaft balancing machine, the spring system supporting the bearing frame part, which is able to oscillate, is made of leaf springs. However, these springs may twist or bulge under load. As a result, the bearing frame has higher resonant frequencies which lie above the rotational speed measurements. The first resonant mode of said higher resonant frequencies is characterised by pitch of the frame upper part and the corresponding resonant frequency is referred to as relaxation resonance. If the balancing rotational speed approaches said relaxation resonance, the imbalance can no longer be recorded precisely. Furthermore, high tension levels are produced due to the excessive increase in resonance in the structure of the machine. This can lead to components failing and could put both the machine and people at risk. The possible rotational speed for measuring imbalance is hence capped by the relaxation resonance. Due to the mass of the tension absorbed by the articulated shaft at the upper part of the bearing frame as well as the mass of the articulated shaft, said relaxation resonance can fall to such an extent that the balancing machine can no longer be operated with the required rotational speed measurement.